JPH05332448A - Speed change control device for hydraulic continuously variable transmission - Google Patents

Abstract

PURPOSE:To prevent the loss of power transmission at the time of backward traveling rotation by a direct connecting clutch for improving the transmission efficiency at the time of forward traveling rotation at a high speed. CONSTITUTION:Angle of inclination of a cam plate 19 for motor is controlled by a governor hydraulic actuator Al, which operates to the vertical standing side in response to a rise of the engine speed, and the inclined angular moment of a plunger 18 for motor. A governor valve G, in which the operating oil pressure rises in response to a rise of the rotating speed of the hydraulic motor side to operate a direct connecting hydraulic clutch C2, is provided. An angle thetar of turn from a neutral position to the backward traveling position of an oil pressure distribution ring 29 is set larger than the angle thetaf of turn from the neutral position to the forward traveling position to enlarge the inclined angular moment under the backward traveling condition larger than that of the forward traveling condition. Even in the case of the maximum of the engine speed, the rotating number of motor never becomes large enough to operate the direct connecting clutch C2.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention comprises a swash plate type hydraulic pump and a swash plate type hydraulic motor which are connected via a hydraulic circuit, and which is provided with a direct coupling hydraulic clutch which operates in accordance with the rotational speed of the hydraulic motor. The present invention relates to a shift control device for a hydraulic continuously variable transmission that performs continuously variable shift control by changing an inclination angle of a swash plate for a motor.

[0002]

2. Description of the Related Art As a hydraulic continuously variable transmission of this type having a direct coupling hydraulic clutch, there is Japanese Examined Patent Publication (Kokoku) No. Sho 42-19728, and when the tilt angle of the swash plate for the motor is substantially upright, that is, the reduction ratio is substantially When the ratio becomes 1: 1, the switching valve interlocked with this operates the direct coupling hydraulic clutch to directly couple the hydraulic motor and the hydraulic pump.

[0003]

DISCLOSURE OF THE INVENTION Problems to be Solved by the Invention
A forward / reverse switching mechanism is not disclosed, but if this conventional technique were applied to reverse rotation, the rotation direction of the hydraulic motor and the rotation direction of the hydraulic pump (input shaft) would be different during reverse rotation of the hydraulic motor. Becomes In such a reverse rotation state, the swash plate for the motor is in the upright state,
When the direct-coupling hydraulic clutch operates, the rotation of the hydraulic motor is braked by the hydraulic pump (input shaft) rotating in the opposite direction to this, resulting in a large loss of power transmission force.

[0004]

SUMMARY OF THE INVENTION It is an object of the present invention to improve the transmission efficiency at the time of forward high speed rotation by providing a direct coupling hydraulic clutch, to perform automatic shift control according to the engine speed and motor load, and to transmit power at the time of reverse rotation. Is the prevention of loss.

[0005]

[Means for Solving the Problems] The swash plate for a motor is designed so that the swash plate for a motor moves to the upright side due to an increase in engine rotation speed, while the load of a hydraulic motor causes the swash plate to move to the tilting side in the opposite direction. The circuit is equipped with a hydraulic distribution ring that can switch the rotation from the neutral position to the forward position and the reverse position.The rotation direction of the hydraulic pump and that of the hydraulic motor are the same when the motor is rotating forward and different when rotating backward. The distribution ring rotation position is set so as to be in the direction, and a governor valve whose operating oil pressure rises according to the increase in the rotation speed on the hydraulic motor side is provided in the oil supply oil passage of the direct coupling hydraulic clutch, and From the neutral position to the reverse position with respect to the rotation angle from the neutral position to the forward position of the hydraulic distribution ring so that the tilting moment with respect to the motor swash plate becomes larger than the tilting moment in the forward moving state. The rotation angle is set to be larger.

[0006]

The swash plate for the motor has an acting force on the upright side by the governor hydraulic actuator in which the working hydraulic pressure rises substantially in proportion to the square of the engine rotating speed, and a pushing force of the motor plunger. The tilt angle is automatically controlled according to the engine rotation speed and the motor load, and the rotation of the hydraulic motor is continuously variable. When the rotational speed of the hydraulic motor increases during forward rotation, the governor valve opens due to centrifugal force, and starts supplying hydraulic oil to the direct coupling hydraulic clutch. Due to the pressure adjusting action of the governor valve, the hydraulic oil pressure supplied from the governor valve rises approximately in proportion to the square of the rotation speed of the hydraulic motor, so the direct coupling hydraulic clutch becomes a half clutch at medium speed. When there is room for kickdown and high speed rotation, the clutch is completely engaged and the gear is in a direct connection. On the other hand, during reverse rotation, the tilting moment of the motor plunger with respect to the motor swash plate is larger than in the forward drive state, so even if the hydraulic pressure of the hydraulic actuator increases due to the increase in engine rotation speed, The swash plate cannot be raised to the upright side, and the motor swash plate is maintained in a state of a large inclination angle (large reduction ratio). Therefore, the direct coupling hydraulic clutch does not operate, and it is possible to prevent the hydraulic motor from being braked by the rotation of the hydraulic pump (input shaft).

[0007]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS FIG. 1 is an overall vertical cross-sectional view (excluding a shift portion) of a hydraulic continuously variable transmission to which the present invention is applied. It is assumed as described in the inside.

First, the overall layout will be described. The front end of the speed-change input shaft 1 is interlocked with an engine (not shown), and on the outer periphery of the input shaft 1, a swash plate type hydraulic motor (axial plunger motor) M and a swash plate type hydraulic pump (axial) are arranged in order from the front side. Type plunger pump) P are arranged in series so as to be aligned in the front-rear direction (axial direction). More specifically, from the front side, the swash plate for motor 1
9, the motor cylinder block 4, the pump swash plate 15, and the pump cylinder block 2 are arranged in series so as to be aligned in the front-rear direction (axial direction).
An intermediate drum 5 is disposed radially outward of the intermediate drum 5 to cover it at intervals, and a direct coupling hydraulic clutch C2 is disposed between the intermediate drum 5 and the pump cylinder block 2. A valve body 3 having a hydraulic circuit is arranged behind the intermediate drum 5 and the pump cylinder block 2.
A rear cover 23 is fixed to the rear side of the valve body 3. Inside the rear cover 23, a hydraulic distribution ring 29 for switching between forward and backward movement, which divides the space portion 24 in the rear cover 23 into an outer space portion 24a and an inner space portion 24b, a distribution ring support shaft 25 for supporting the same, and a main ring. A spool 30 for switching the clutch and the like are arranged. A hydraulic actuator A1 for controlling the inclination angle of the motor swash plate 19 to perform automatic shift control is disposed above and in front of the motor swash plate 19.

The outline of the hydraulic pump P will be described. The pump cylinder block 2 has an inner peripheral side spline-fitted to the outer periphery of the rear end portion of the input shaft 1 and rotates integrally with the input shaft 1. The pump cylinder block 2 has an odd number of, for example, five cylindrical holes 71 formed at equal intervals in the circumferential direction. Each cylindrical hole 71 is formed parallel to the input shaft 1 and It has an opening toward. Each cylindrical hole 7
1, a bottomed cylindrical pump plunger 14 is fitted so as to be slidable in the axial direction so as to project forward, and a gentle spherical shape portion 14a is formed at the front end of each pump plunger 14. And is in contact with the pump swash plate 15 on the front side. The pump swash plate 15 serves as a guide for the reciprocating motion of the pump plunger 14, and the thrust bearing 16
Through the rear end slope 4a of the motor cylinder block 4
It is installed in. A coil spring 17 for biasing the pump plunger 14 forward is contracted between each of the pump plungers 14 and the rear end surface of the cylindrical hole, whereby the pump swash plate 15 is pressed and the pump swash plate 15 is pressed. In addition to preventing falling, the pump cylinder block 2 is pressed against the rear valve body 3 to improve the sealing performance of the sliding seal surface 2b during low rotation, and further the pump plunger 14 is urged forward. As a result, the self-suction capability is imparted to the hydraulic pump P.

The outline of the hydraulic motor M will be described. The motor cylinder block 4 is integrally fastened to the valve body 3 via the intermediate drum 5, and the inner peripheral side of the cylinder block 4 is rotatably fitted to the input shaft 1 with the annular oil passage 101 interposed therebetween. The case 22 is supported by bearings 72 and the like. The motor cylinder block 4 is arranged at equal intervals in the circumferential direction and has an odd number of, for example, nine cylindrical holes 73.
The cylindrical hole 73 is formed parallel to the input shaft 1 and opens forward as in the case of the hydraulic pump P. A cylindrical motor plunger 18 having a bottom is fitted into each cylindrical hole 73 so as to be slidable in the axial direction and capable of projecting forward, and the front end of each motor plunger 18 has a gentle spherical shape. 18a is formed and is in contact with the front motor swash plate 19. The motor swash plate 19 serves as a guide for the reciprocating motion of the motor plunger 18, and is supported by the swash plate holder 21 on the front side via the rolling elements 20. Plunger 1 for each motor
A coil spring 37 for urging the coil 8 forward is contracted between the rear end face of the cylindrical hole and the self-suction force is applied to the hydraulic motor M.

The oil passage structure will be described. In FIG. 1, a front end portion of an annular oil passage 101 formed on the outer periphery of the input shaft 1 is connected to an unillustrated HST charging feed pump or the like via an annular oil chamber 102, and is connected to an annular oil chamber 102 from an oil pan or the like. The hydraulic oil is supplied to the annular oil passage 101 via the. At the rear end of the annular oil passage 101, there is a direct-connection hydraulic clutch cooling oil passage 103 that extends through the input shaft 1 to the direct-connection hydraulic clutch C2, the motor cylinder block 4, the intermediate drum 5, and the valve body. Hydraulic fluid chamber 105 of the hydraulic clutch C2 for direct connection
To the outer space 24a in the rear cover 23 through the direct coupling hydraulic clutch operating oil supply oil passage 106, the motor cylinder block 4, the intermediate drum 5 and the valve body 3 as shown in FIG. Oil passage 108 communicates.

A check valve 74 that opens to the outer space 24a is provided at the outlet of the charging oil passage 108 to the outer space 24a.
Is provided with a valve seat portion 75 on the valve body 3 and a steel ball 45 seated on the seat portion 75 is incorporated. Further, the check valve 74 tilts the valve seat portion 75 so that the centrifugal force generated in the steel ball 45 by the rotation of the valve body 3 acts on the steel ball 45 backward (in the valve opening direction). .. A branch oil passage 108a reaching the inner space portion 24b is formed in the charging oil passage 108, and the outlet of the branch oil passage 108a to the inner space portion 24b is opened toward the inner space portion 24b. A check valve 77 is provided. The check valve 77 includes a valve seat 46, a compression coil spring 48, a steel ball 49 that is seated on the valve seat 46 by the elastic force of the compression coil spring 48, and the pressing rod 4.
7 and the like, and is arranged at the approximate center of the valve body 3.

An outline of an oil passage in which the working oil supplied from the above-mentioned charging oil passage 108 to the space portion 24 reaches the hydraulic motor M via the hydraulic pump P of FIG. 1 and returns to the space portion 24 again will be described. .. Regarding the hydraulic pump P, the valve body 3 communicates with the pump cylindrical hole 71 in the suction stroke and the outer space portion 24a as shown in the upper side of FIG. To the pump-side suction oil passage 110 for sucking the working oil, the pump cylindrical hole 71 in the discharge stroke and the inner space portion 24b are communicated with each other through the pump cylindrical hole 71 inward as shown in the lower side. A pump-side discharge oil passage 111 for discharging hydraulic oil is formed in the space 24b. Regarding the hydraulic motor M, the valve body 3,
In the intermediate drum 5 and the motor cylinder block 4, there are nine motor-side oil passages 11 communicating with the motor cylindrical holes 73.
3, the motor-side oil passage 113 communicates with the space 24 and the motor cylindrical holes 73, but the rear end port of the motor-side oil passage 113 has the same circle centered on the rotation axis. By the rotation of the valve body 3 which is arranged on the circumference and rotates integrally with the motor cylinder block 4,
As shown in the lower side of FIG.
The hydraulic oil is forced into the motor cylindrical hole 73 from the inner space 24b by communicating with the inner space 24b. On the other hand, as shown in the upper side, the motor cylindrical hole 73 in the discharge stroke is formed in the outer space 2b.
4a in communication with the motor cylindrical hole 73 to the outer space 24
The hydraulic oil is discharged to a.

The swash plate holder 21 shown in FIG. 1 has a transmission case 22 via a trunnion shaft 22a which is perpendicular to the input shaft 1.
Is supported on the trunnion shaft center and can be tilted about the center of rotation, and the tilting moment caused by the hydraulic actuator A1 for automatic shift control and a motor plunger 18 described later causes tilting with respect to a plane perpendicular to the input shaft center. Change. Thereby, the stroke amount of the motor plunger 18 can be adjusted to continuously change the motor capacity. The gear ratio of the hydraulic motor M to the hydraulic pump P is given by the following equation. Gear ratio = Pump speed / Motor speed = 1 + Motor capacity / Pump capacity As is clear from the above equation, if the motor capacity is changed from 0 to a certain value, the gear ratio can be changed from 1 to a necessary value. it can. The motor capacity is the plunger 1 for the motor.
When the motor swash plate 19 is in a posture perpendicular to the input shaft center, the stroke becomes 0, the motor capacity becomes 0, and the gear ratio is 1. Then, as the motor is inclined as shown in FIG. 1, the gear ratio is increased as the motor capacity is increased.

A detailed structure for controlling the inclination angle of the swash plate holder 21 will be described. Hydraulic actuator A1
Has a piston 35 having a spherical sliding portion slidably fitted in a cylinder 36 formed integrally with the transmission case 22, and the piston 35 is connected to a rear end spring bearing ring of the cylinder 36. Coil spring 81 compressed between
Is urged forward. A rod is integrally formed with the piston 35, and a rod end 34 protruding rearward is connected to the piston 35.
4 is a swash plate holder 21 for a motor via a ball joint mechanism.
Is connected to the upper end of. Oil chamber 82 in the cylinder 36
The transmission case 2 through the throttle 39 for adjusting the flow rate.
It communicates with the oil passage 120 in the wall of No. 2. The oil passage 120
Is connected to a so-called governor hydraulic power source whose pressure changes substantially in proportion to the square of the rotation speed of the engine, and hydraulic oil having the above pressure is supplied. Therefore, the operating oil pressure of the oil chamber 82 of the cylinder 36 rises in response to the increase of the engine rotation speed (approximately proportional to the square of the engine rotation speed), and the coil spring 8
Against 1, the motor swash plate 19 is raised to a state perpendicular to the input shaft 1.

On the other hand, the trunnion shaft 2 of the swash plate holder 21
The position of 2a is eccentric to the side opposite to the hydraulic actuator side (lower side in the figure) with respect to the motor rotation axis, and
It is set to the rear side of the motor swash plate 19, so that the forward load of the motor plunger 18 causes the motor swash plate 19 to tilt relative to the motor swash plate 19 in the T1 direction to increase the reduction ratio. A tilting moment acts. The tilting moment increases as the motor load increases and the operating hydraulic pressure in the motor plunger 18 increases, and the tilting of the motor swash plate 19 increases to increase the gear ratio (reduction ratio). To do.
That is, as described above, the shift control device operates as a governor hydraulic pressure in the cylinder oil chamber 82 that acts to raise the motor swash plate 19 in response to an increase in the engine rotation speed (substantially proportional to the square of the engine rotation speed). A motor swash plate 19 that is eccentrically arranged so as to incline by the pressing load (reaction force) of the motor plunger 18 corresponding to an increase in motor load, and a coil spring 81 that urges the piston 35 in a direction in which the inclination angle increases. It is possible to automatically control the gear ratio corresponding to the engine rotation speed and the motor load by balancing these three elements.

The structure inside the rear cover 23 will be described. As described above, the rear cover 23 is integrally fastened to the rear end surface of the valve body 3 so as to rotate integrally with the valve body 3.
The space portion 24 is formed between the inner surfaces of the. The hydraulic distribution ring support shaft 25 arranged in the space 24 is arranged substantially coaxially with the input shaft 1, and the radial bearing 2 is provided.
It is rotatably supported by the rear cover 23 via 7, 28, and a thrust bearing 26 is arranged between the rear cover 23 and the rear side step portion. It regulates movement. Support shaft 25
A cylindrical portion 25a that is eccentric from the support axis is integrally formed at the front end of the, and the bottomed cylindrical hydraulic distribution ring 29 is fitted to the outer peripheral surface of the eccentric cylindrical portion 25a. The space 24 in the rear cover 3 is connected to the distribution ring 29 as described above.
The outer space portion 24a on the outer side and the inner space portion 2 on the inner side
It is divided into 4b. The front end surface of the hydraulic distribution ring 29 is in sliding contact by being pushed out to the rear end surface of the valve body 3 by the hydraulic oil pressure in the inner space portion 24b, and the pressing strength automatically changes due to the pressure. It has a gap compensation function. By rotating the support shaft 25, the hydraulic distribution ring 29 can be switched between the forward drive position F, the neutral position N, and the reverse drive position R as shown in FIG. There is.

In FIG. 4, which is a sectional view taken along line IV-IV of FIG. 3, a line 22'a is parallel to the trunnion shaft 22a,
A line that passes through the motor axis and is perpendicular to this line 22'a (hereinafter, tentatively referred to as "forward / backward separation line") L,
There is a top dead center (most contraction point) U1 and a bottom dead center (most extension point) U2 of the motor plunger 18. When the X1 direction is the forward rotation direction, the neutral position N of the distribution ring 29 is set at a position where the center of the distribution ring 29 passes along the forward / backward separation line L and is most biased to the top dead center U1 side. There is. The forward drive position F is set to a position rotated by an angle θf (approximately 90 °) from the neutral position N to the motor forward rotation direction X1 side around the support axis, but the reverse drive position R is set from the neutral position N. It is set at a position rotated by an angle θr larger than the angle θf in the motor reverse rotation direction X2 around the support axis, and the difference between both angles (θr-θf) = θa is, for example, approximately 5 °. It is set. That is, Kn is the center of the distribution ring in the neutral position, Kf is the center of the distribution ring in the forward position,
If Kr is the center of the distribution ring in the reverse position, in the illustrated embodiment, the distribution ring center Kf in the forward direction is approximately the above line 22 '.
The center of the distribution ring Kr when the vehicle is moving backward is displaced toward the bottom dead center U2 side from the line 22'a, and the line 22'a
The displacement angle θa with respect to is approximately 5 °.

Of the pump-side oil passages 110 and 111, there are two pump-side suction oil passages 110 communicating with the outer space 24a.
The rear end ports are individually formed, and the rear end ports are located at the upper outer end portion, and the outer space portion 24 is always maintained even when the distribution ring 29 is in the forward position F, the neutral position N, or the rearward position R.
It communicates with a. Three pump-side discharge oil passages 111 communicating with the inner space portion 24b are formed, their rear end ports are located in the lower portion near the axis, and the distribution ring 29 is in the forward movement position F and the neutral position N. Alternatively, in any state of the reverse drive position R, it is always in communication with the inner space portion 24b.

On the other hand, the rear end ports of the nine motor-side oil passages 113 are arranged at equal positions in the radial direction at equal intervals on the same circumference. The range from the bottom dead center U2 to the top dead center U1 around the forward rotation direction X1 (the right side range in the figure) is W1, the top dead center U
The range from 1 to the bottom dead center U2 around the forward rotation direction X1 (left side range in the figure) is W2. When the distribution ring 29 is at the neutral position N, the motor-side oil passage 113, which is approximately within the range of the bottom dead center U2 side of the line 22'a, communicates with the outer space portion 24a.
The motor side oil passage 113 in the range on the side of the top dead center U1 communicates with the inner space portion 24b, and the hydraulic motor M does not rotate. When the distribution ring 29 is at the forward position F, the motor side oil passage 113 in the range W1 communicates with the outer space portion 24a, and the motor side oil passage 113 in the range W2 communicates with the inner space portion 24b. That is, during forward rotation, the motor side oil passage 113 in the range W1 communicates with the motor cylindrical hole 73 in the discharge stroke, and the motor side oil passage 113 in the range W2 extends in the expansion stroke motor cylindrical hole 73. And the hydraulic motor M rotates in the forward rotation direction X1. When the distribution ring 29 is at the reverse position R, the motor-side oil passage 113 within the range W'1 deviated from the range W1 by the angle θa toward the X2 direction side.
Is communicated with the inner space portion 24b, and the motor side oil passage 113 in the range W'2 deviated from the range W2 in the X2 direction by the angle θa is communicated with the outer space portion 24a. That is, in reverse, the motor-side oil passage 113 in the range W'1 communicates with the motor cylindrical hole 73 in the expansion stroke, and the motor-side oil passage 113 in the range W'2 for the discharge stroke motor. It communicates with the cylindrical hole 73 and rotates the hydraulic motor M in the reverse rotation direction X2.
In this case, the virtual center of the pressing load of the motor plunger 18 is closer to the bottom dead center U2 side (upper side) than in forward rotation.
Therefore, the plunger pressing load on the lower side (top dead center U1 side) of the trunnion shaft 22a is reduced by θa / 180, and the upper side of the trunnion shaft 22a (bottom dead center U2 The plunger pressing load (on the side) increases by θa / 180, and as a whole, the tilting moment increases as compared with the tilting moment during forward movement. In other words, since the rotation angle θr at the reverse position is set larger than the rotation angle θf at the forward position and is displaced toward the bottom dead center U2 side (upper side), the pressing load of the motor plunger 18 at the time of reverse movement is increased. The imaginary center of is moved higher than when moving forward, and the tilting moment with respect to the motor swash plate 19 is increased compared to when moving forward.
Therefore, the force for tilting the motor swash plate 19 becomes large. In this embodiment, the hydraulic actuator A1 of FIG.
By adjusting the load of the spring 81 of the motor, even if the engine speed becomes maximum during reverse rotation, the motor plunger 1
The displacement angle θa is set so that the force for tilting the motor swash plate 19 due to the tilting moment of 8 is larger than the force for raising the motor swash plate 19 to the upright side by the hydraulic actuator A1. And during reverse rotation,
The motor swash plate 19 is always kept at the maximum tilt position (maximum gear ratio position), whereby the hydraulic motor M rotates in the low speed range, and the governor valve G that operates in association with the rotation of the hydraulic motor M opens. Therefore, the direct connection hydraulic clutch C2 is not activated.

The front end surface (sliding contact surface) of the distribution ring 29 has a pair of wedge-shaped grooves 7 on each of the inner peripheral end side and the outer peripheral end side.
8a and 78b are formed. The formation positions of the grooves 78a and 78b are located near the top dead center U1 and the bottom dead center U2 when the distribution ring 29 is located at the forward drive position F or the reverse drive position R, and both the points U1 and U2. On the other hand, they are formed at positions projecting from the inside and the outside, respectively.
As a result, when the motor-side oil passages 113 pass through the top dead center U1 and the bottom dead center U2, the inner space and the outer space are gradually closed from the distribution ring 29 and then gradually opened to the other space. It is supposed to do.

The structure of the direct coupling hydraulic clutch C2 will be described. In FIG. 1, the direct coupling hydraulic clutch C2 is a wet multi-plate type clutch, and includes a large number of friction plates 6.
And a large number of separator plates 7 arranged alternately with the friction plates 6, and the friction plates 6 are axially slidably engaged with splines 2a formed on the outer periphery of the pump cylinder block 2. do it,
It is designed to rotate integrally with the cylinder block 2. A pair of pressure plates 9 capable of sandwiching the plates 6 and 7 in the axial direction are arranged at both axial ends of the direct coupling hydraulic clutch C2, and the pressure plates 9 are contracted between them. The pressure plates 9 are urged in a direction away from each other in the axial direction, that is, the clutch-off side by a coil spring (not shown).

The hydraulic piston 11 for actuating the direct coupling hydraulic clutch C2 is fitted in the annular recess of the valve body 3 so as to be slidable in the axial direction, and projects forward so that both plates 6, 6 are interposed between the pressure plates 9. The hydraulic motor M and the hydraulic pump P can be turned on (directly connected) by pressing the 7 into contact with each other.

The structure of the centrifugal governor valve G of the direct coupling hydraulic clutch C2 will be described with reference to FIG. The centrifugal governor valve G is provided in the motor cylinder block 4 in the middle of the direct coupling hydraulic clutch operating oil supply oil passage 106, and is composed of a stepped moving spool 40, a fixed sleeve 41, and the like. The fixed sleeve 41 is fixed to a hole 84 formed in the motor cylinder block 4, and the stepped moving spool 40 is fitted in the sleeve 41 so as to be movable in the radial direction. The spool 40 is biased radially inward (closed position).
Reference numeral 121 denotes a drain hole which communicates with the outside when the transmission hole 22 is closed to prevent leakage of oil from the direct coupling hydraulic clutch side.
It is designed to be returned inside. The motor cylinder block 4 rotates, and the stepped moving spool 40 is rotated by centrifugal force.
Moves outwards against the spring 42, and when the rotational speed reaches a predetermined value, the inlet side passage 122 and the outlet side passage 123 communicate with each other, and hydraulic oil flows into the annular step surface portion 40a of the spool 40. .. Since the spool 40 has an area difference (area of the step surface portion) between the small diameter portion and the large diameter portion, the hydraulic oil in the sleeve 41 generates a force that pushes the spool 40 inward in the radial direction. Due to the balance of the inward pushing force due to the above area difference and the outward force due to the centrifugal force, a hydraulic oil pressure approximately proportional to the square of the rotation speed is generated, and the direct coupling hydraulic clutch C2
Is supplied to the oil chamber 105. The hydraulic oil pressure to the direct coupling hydraulic clutch C2 increases as the rotational speed increases, whereby the hydraulic oil is controlled so as to go from the half-clutch state at medium speed to the complete clutching state at high speed. ..

Referring to FIG. 3, the structure of the main clutch C1 for connecting and disconnecting the flow of hydraulic oil between the hydraulic pump P and the hydraulic motor M will be described. The main clutch spool 30 is the support shaft 2
5 is rotatably fitted in the inner peripheral hole of the support shaft 25, and the thrust load balls 31, 32
It is supported so that it can smoothly rotate even with respect to an axial load. At the front end of the spool 30, a vertical oil passage 115 that opens to the inner space 24b and extends rearward along the spool axis, and an oil passage 116 that extends radially outward from the rear end of the vertical oil passage 115. Is formed, while the support shaft 25 has
An oil passage 117 extending outward and opening to the outer space portion 24a is formed at the same axial position as the oil passage 116, and the rotation of the spool 30 causes the oil passages 116, 11 to move.
The main clutch C1 is turned on and off by disconnecting and connecting the seven.

The operation of the entire hydraulic transmission and the operation related to the present invention will be described. When the pump cylinder block 2 is rotated by the rotation of the input shaft 1, the high-pressure hydraulic oil flows from the cylindrical hole 71 accommodating the pump plunger 14 in the discharge stroke to the inner space 24b through the pump-side discharge oil passage 111. Is discharged, and the high-pressure hydraulic oil passes through a part of the motor-side oil passage 113 and is press-fitted into the cylindrical hole 73 that houses the motor plunger 18 in the expansion stroke. On the other hand, the hydraulic oil discharged from the cylindrical hole 73 accommodating the motor plunger 18 in the discharge stroke to the outer space 24a through the remaining motor side oil passage 113 passes through the pump side suction oil passage 110. It is returned to the pump cylindrical hole 71 having the pump plunger 14 in the suction stroke. Meanwhile, the sum of the reaction torque given to the motor cylinder block 4 by the pump plunger 14 in the discharge stroke via the pump swash plate 15 and the reaction torque received by the motor plunger 18 in the expansion stroke from the motor swash plate 19 The motor cylinder block 4 is rotated. The swash plate 19 for the motor is operated by the governor hydraulic actuator A1 whose operating oil pressure rises substantially in proportion to the square of the engine rotation speed, and the tilting moment due to the pressing load of the motor plunger. And the tilt angle is controlled according to the motor load, and thereby continuously variable transmission control is performed.

When the hydraulic motor M is stopped, the governor valve G is closed because the centrifugal force is not applied. That is, the oil chamber 105 of the hydraulic clutch C2 for direct connection
The supply of hydraulic oil to is shut off. When the rotation speed of the hydraulic motor M increases during forward rotation, the governor valve G opens due to centrifugal force and starts supplying hydraulic oil to the direct coupling hydraulic clutch C2. Due to the pressure adjusting action of the governor valve G, the hydraulic oil pressure supplied from the governor valve G rises approximately in proportion to the square of the rotation speed of the hydraulic motor. Therefore, the direct coupling hydraulic clutch C2 becomes a half clutch at medium speed. When there is room for kickdown and high-speed rotation, it is completely clutched and is in a direct connection state. On the other hand, at the time of reverse rotation, the tilting moment of the motor swash plate 19 toward the reduction ratio increasing side by the motor plunger 18 becomes larger than that at the time of forward movement, so that the hydraulic actuator A1 operates due to the increase of the engine speed. Even if the hydraulic pressure increases, the motor swash plate 19 cannot be raised to the upright side, and the motor swash plate 19 is maintained at the maximum inclination angle (maximum reduction ratio). Therefore, the direct coupling hydraulic clutch C2 does not operate, and the hydraulic pump P
Can prevent the rotation of the hydraulic motor M from braking the rotation of the hydraulic motor M.

[0028]

[Another embodiment] (1) Reverse rotation angle θr of distribution ring 29
Is a tilting moment that keeps at least the motor rotation below the direct coupling clutching point, even when the tilting moment when moving backward is larger than the tilting moment when moving forward and the engine speed when moving backward is maximum. Therefore, for example, when the forward rotation angle θf is set to be smaller than 90 °,
θr is larger than θf, but may be smaller than 90 °.

[0029]

As described above, according to the present invention, the acting force on the upright side by the governor hydraulic actuator A1 and the tilting moment due to the pressing force of the motor plunger enable the engine speed and the motor load to be changed according to the engine rotation speed and the motor load.
Continuously variable speed control is possible. At the time of forward rotation, the hydraulic clutch C2 for direct connection will automatically be operated at a predetermined motor rotation speed or higher.
To automatically improve the power transmission efficiency during high-speed rotation. On the other hand, during reverse rotation, when the hydraulic pump P and the hydraulic motor M are rotating in different directions, how much engine speed Can be maintained in a state of a large inclination angle even if the
As a result, the operation of the direct coupling hydraulic clutch C2 can be prevented, and the loss of power transmission during reverse rotation can be prevented.

[Brief description of drawings]

FIG. 1 is an overall vertical cross-sectional view of a hydraulic continuously variable transmission to which the invention of the present application is applied.

FIG. 2 is an enlarged vertical cross-sectional partial view of a charging oil passage circuit portion of the same hydraulic transmission as in FIG.

FIG. 3 is an enlarged partial view of a portion III in FIG.

FIG. 4 is a sectional view taken along line IV-IV in FIG.

FIG. 5 is an enlarged vertical sectional view of the governor valve.

[Explanation of symbols]

 1 Input Shaft 2 Pump Cylinder Block 4 Motor Cylinder Block 15 Pump Swash Plate 18 Motor Plunger 19 Motor Swash Plate 29 Hydraulic Distribution Ring A1 Hydraulic Actuator P Swash Plate Hydraulic Pump M Swash Plate Hydraulic Motor C2 Direct Connection Hydraulic Clutch G Governor valve

Claims (1)

[Claims] 1. A swash plate type hydraulic pump and a swash plate type hydraulic motor are connected via a hydraulic circuit, and a direct connection hydraulic clutch that operates in accordance with the rotational speed of the hydraulic motor side is provided. In a hydraulic continuously variable transmission that performs continuously variable shift control by changing the tilt angle, the swash plate for the motor has the swash plate for the motor moved to the upright side as the engine rotation speed increases.
On the other hand, the load of the hydraulic motor causes it to move in the opposite direction to the tilting side, and the hydraulic circuit is provided with a hydraulic distribution ring capable of rotating and switching from the neutral position to the forward position and the reverse position. Set the distribution ring rotation position so that the rotation direction of the hydraulic motor and that of the hydraulic motor are the same when the motor is rotating forward and are different when rotating backward, and the oil supply oil passage of the direct coupling hydraulic clutch is connected to the hydraulic motor side. A governor valve whose operating oil pressure rises in accordance with the increase in the rotation speed of is installed, and the neutral pressure of the hydraulic distribution ring is adjusted so that the tilting moment with respect to the motor swash plate in the reverse drive state is larger than the tilting moment in the forward drive state. For the rotation angle from the position to the forward position,
A shift control device for a hydraulic continuously variable transmission, characterized in that a large turning angle is set from a neutral position to a reverse position.